JPH07122441B2 - Brake valve - Google Patents

Description

Description: TECHNICAL FIELD The present invention relates to a brake valve suitable for use in a hydraulic circuit that drives a hydraulic motor of, for example, a hydraulic excavator.

Technical Background Generally, construction machinery such as hydraulic excavators and hydraulic cranes are
Driven by a hydraulic motor as a hydraulic actuator,
Although the vehicle is driven to swing, it is subjected to excessive inertial load when traveling or stopping swinging, so a brake valve is installed in the hydraulic circuit, and the brake valve converts the inertial load into thermal energy of the pressurized oil and absorbs it.

This brake valve is composed of a counterbalance valve provided between the hydraulic power source and the hydraulic actuator, a pair of relief valves provided closer to the actuator than the counterbalance valve, and a pair of check valves provided between each of the relief valves and the counterbalance valve (see, for example, Japanese Utility Model Publication No. 62-31681).

6 to 9, the above-mentioned prior art brake valve will be shown and explained.

In the figure, reference numeral 1 denotes a hydraulic motor as an actuator, and the hydraulic motor 1 includes a control pin 8, a rotor, and a
The hydraulic motor 1 is configured as a radial piston type hydraulic motor, comprising a piston 11, a piston 14, etc. The hydraulic motor 1 has its supply and discharge ports connected to a hydraulic pump 3 as a hydraulic source via a pair of pipes 2A, 2B, etc., and the hydraulic pump 3 is connected to a tank 4.
It is driven to rotate by the pressure oil sucked in and discharged from the

Reference numeral 5 denotes a directional control valve located between the hydraulic motor 1 and the hydraulic pump 3 and provided in the middle of each of the paths 2A and 2B. The directional control valve 5 is manually operated by the driver to:
At the neutral position (a), the hydraulic motor 1 is stopped, and at the switching positions (b) and (c), the hydraulic motor 1 is rotated in the directions of the arrows F and R.

Reference numeral 6 denotes a brake valve located between the hydraulic motor 1 and the directional control valve 5 and provided in the middle of each of the pipelines 2A and 2B. The brake valve 6 is connected to a counterbalance valve 23 and a relief valve 24, which will be described later.
The brake valve 6 is composed of relief valves 28A, 22B, check valves 39A, 39B, etc., and is provided integrally with the hydraulic motor 1. When the directional control valve 5 is switched to the neutral position (a), one of the relief valves 28A, 28B opens to relieve the pressure oil on the high-pressure side of each of the pipes 2A, 2B to the low-pressure side,
It applies a braking force to the hydraulic motor 1.

Next, the motor 1 will be described with reference to FIG.

In the figure, reference numeral 7 denotes a motor casing in the form of a cylindrical cover. A cam surface 7A is formed on the inner periphery of the motor casing 7 over the entire periphery, and the opening is covered by a valve casing 18 described below.

Reference numeral 8 denotes a control pin that protrudes from approximately the center of a valve casing 18 and is formed integrally with the valve casing 18, with its axis offset by a predetermined distance from the axis of the cam surface 7A. As shown in FIG. 8, a pair of supply and discharge passages 9, 10 that form supply and discharge ports for the hydraulic motor 1 are formed in the axially intermediate portion of the control pin 8 and open onto the outer peripheral surface.

11 denotes a rotor rotatably mounted on the control pin 8;
The rotor 11 has a plurality of cylinders 11A (only one shown) formed radially at a predetermined angular interval in the circumferential direction, and the cylinders 11A intermittently pass through the supply and discharge passages 9, 10 via ports 11B. The axis of the rotor 11 is eccentric with respect to the axis of the cam surface 7A, so a crescent-shaped eccentric space 12 is formed between the outer periphery of the rotor 11 and the cam surface 7A.

Reference numeral 13 denotes an output shaft integrally provided on one end surface of the rotor 11, and the output shaft 13 is connected to an external inertial body via a reduction gear (not shown) or the like.
rotates integrally with the rotor 11,
It outputs the rotational force of 11 to the outside.

Reference numeral 14 denotes a piston reciprocally provided in each cylinder 11A of the rotor 11, 15 denotes a ball swingably provided in the piston 14, and 16 denotes shoes positioned between the ball 15 and the cam surface 7A and spaced apart in the circumferential direction.
The shoe 16 has its tip end fitted onto the ball 15 and its base end slidably attached to the cam surface 7A via guides 17, 17.

Next, the brake valve 6 will be described with reference to FIGS.

In the figure, reference numeral 18 denotes a valve casing that covers the opening side of the motor casing 7 and is firmly fixed to the motor casing 7 with a plurality of bolts 19, 19, .... Oil passages 20A and 20B, which will be described later, are integrally formed in the valve casing 18, and a counterbalance valve 23, relief valves 28A and 28B, check valves 39A and 38B, etc. are also provided.

Reference numerals 20A and 20B denote oil passages provided in the valve casing 18 and constituting part of the pipelines 2A and 2B. The oil passages 20A and 20B are oil passages on the hydraulic source side connected to the hydraulic pump 3 via the directional control valve 5.
The hydraulic power source side oil passages 21A, 21B and actuator side oil passages 22A, 22B are connected to the hydraulic motor 1 via the supply and discharge passages 9, 10. The hydraulic power source side oil passages 21A, 21B and the actuator side oil passages 22A, 22B pass through check valves 39A, 32B.

Reference numeral 23 denotes a counterbalance valve provided in the valve casing 18, positioned closer to the hydraulic pump 3. The counterbalance valve 23 is generally composed of a spool slide hole 24 integrally formed in the valve casing 18, and a spool 25 provided slidably in the spool slide hole 24. The spool 25 is provided with a land 25A for connecting and disconnecting the oil passages 21A and 22A, and a land 25B for connecting and disconnecting the oil passages 21B and 22B. The spool 25 has oil chambers 26A and 26B at both ends.
26B, and return springs 27A and 27B are inserted into the oil chambers 26A and 26B.
The counterbalance valve 23 operates in conjunction with the directional control valve 5 and is switched from the neutral position (a) to the switching positions (b) and (c).

28A and 28B are located near the hydraulic motor 1 side and are connected to the valve casing 1
9, the pair of relief valves 28A and 28B are connected to a valve guide 29 (described later).
A crossover relief valve with a shockless function is configured from the main valve bodies 32A and 32B, the pistons 36A and 36B, and the like.

Reference numerals 29A and 29B denote valve guides whose tip ends face into the respective actuator-side oil passages 22A and 22B and form part of the respective relief valves 28A and 28B, and flow holes 30A and 30B are formed at the tip ends of the respective valve guides 29A and 29B.

Reference numerals 31A and 31B denote cylindrical valve seat members positioned opposite the valve guides 29A and 29B and communicating between the actuator-side oil passages 22A and 22B. Reference numerals 32A and 32B denote main valve bodies slidably mounted on the valve guides 29A and 29B so as to be seated on and separated from the valve seat members 31A and 31B. The main valve bodies 32A and 32B are supported by valve springs 33A and 33B.
The main valve bodies 32A, 32B are normally biased in the valve closing direction by the respective throttle passages 34A, 34B formed in the axial direction.

Here, for each of the main valve bodies 32A and 32B, the diameter of the seating portion between the valve seat members 31A and 31B is d1, and the diameter of the sliding portion between the valve guides 29A and 29B is d
If 2, then the relationship d1>d2 is true.

The pressure P1 in each of the valve seat members 31A and 31B is
When the pressure P1 is equal to the pressure P2 in A, 29B, when the pressure P1 reaches the valve opening set pressure P0, the main valve bodies 32A, 32B are lifted off by the valve seat members 31A, 31B against the valve springs 33A, 33B due to the difference in pressure receiving area determined by the diameters d1, d2, and the valves are opened to connect the oil passages 22A and 22B, thereby performing a high pressure relief function.

On the other hand, the accumulators 38A and 38B, which will be described later, are activated.
When the pressure P2 in each of the valve guides 29A, 29B is kept lower than the pressure P1 in each of the valve seat members 31A, 31B, even if the pressure P1 is lower than the valve opening set pressure P0, the main valve bodies 32A, 32B are closed by the valve springs 3
3A, 33B, and the oil passages 22A, 2
The valve opens to allow passage between 2B, providing low pressure relief.

35A and 35B are the outer periphery of each valve guide 29A and 29B and the separate casing 18
The inner end of the actuator-side oil passage 22 is formed between the actuator-side oil passage 22 and the inner end of the actuator-side oil passage 22.
The pistons 36A and 36B are cylindrical oil chambers that communicate with the valve guides 29A and 29B, respectively, and are cylindrical pistons that are slidably provided in the oil chambers 35A and 35B. The pistons 36A and 36B are connected to throttle passages 37A and 37B that are formed in the axially intermediate portions of the valve guides 29A and 29B, respectively, and to the oil chambers 35A and 35B.
5B, as shown in FIG. 6, form accumulators 38A and 38B.
The pressure oil in the valve guides 29A and 29B passes through the throttle passages 34A and 34B.
and then flows into the oil chambers 35A and 35B via the throttle passages 37A and 37B.
When the oil flows into the outer end space of the piston 36A, 36B,
The pistons 36A, 36B move toward the tip ends of the valve guides 29A, 29B until they abut against the valve casing 18, and during this time, the relief valves 28A, 28B maintain their valve opening set pressures low. The period during which the pistons 36A, 36B are moving corresponds to the constant pressure relief time.

39A, 39B are the counterbalance valve 23 and the relief valves 28A, 28B
The check valves 36A and 39B are located between the oil passages 20A and 20B, and are provided in the middle of the oil passages 20A and 20B.
Valve springs 40A, 40B are provided between the hydraulic power source side passages 21A, 21B and the actuator side passages 22A, 22B, respectively, and valve seats 41A, 41B are provided between the hydraulic power source side passages 21A, 21B and the actuator side passages 22A, 22B.
When pressure oil from the hydraulic pump 3 flows into each of the pressure-source-side oil passages 21A, 21B, the check valves 39A, 39B open against the spring force of the valve springs 40A, 40B, allowing the pressure oil to flow toward each of the actuator-side oil passages 22A, 22B while preventing the pressure oil from flowing in the opposite direction.

42 is provided in the valve casing 18 at a position closer to the hydraulic pump 3 side than the counterbalance valve 23, and each hydraulic source side oil passage
21A, 21B, and the shuttle valve 4
Reference numeral 2 denotes a valve for selecting the high-pressure oil passage from among the hydraulic pressure source oil passages 21A, 21B, and supplying a portion of the pressure oil as pilot pressure to a brake device (not shown) or the like.

The brake valve according to the prior art has the above-mentioned structure, and its operation will now be described.

First, when the driver switches the directional control valve 5 from the neutral position (a) to the switching position (b), the pressure oil discharged from the hydraulic pump 3 flows into the hydraulic source side oil passage 21A via the pipe 2A.
The check valve 39A opens against the valve spring 40A due to the pressure difference between the hydraulic pressure source side oil passage 21A and the actuator side oil passage 22A, and the pressure oil in the hydraulic pressure source side oil passage 21A flows into the cylinder 1 of the rotor 11 through the actuator side oil passage 22A and each supply/discharge passage 9, etc.
It flows into 1A.

As a result, when the piston 14 reciprocates within the cylinder 11A, the reciprocating movement of the piston 14 causes the shoe 16 to slide along the cam surface 7A, and the rotor 11 rotates around the control pin 8. The rotational force of the rotor 11 is then output to the outside from the output shaft 13, rotating the inertial body.

In addition, when pressure oil flows into the hydraulic source side oil passage 21A,
A part of this pressure oil flows through the throttle passage of the spool 25 into the oil chamber 26
8. As a result, the land 25B also moves rightward, connecting the hydraulic pressure source side oil passage 21B and the actuator side oil passage 22B. In other words, the counterbalance valve 23 switches from the neutral position (a) to the switching position (b) shown in FIG. 6. Then, the pressure oil pushed out from the cylinder 11A by the compression stroke of the piston 14 flows into the actuator side oil passage 22B via the supply and discharge passages 10, etc.
B, and is discharged to the outside of the brake valve 6 via the actuator side oil passage 22B and the hydraulic pressure source side oil passage 21B, and further returns to the tank 4 via the directional control valve 5 and the pipe 2B.

On the other hand, when the driver switches the directional control valve 5 from the switching position (b) to the neutral position (a) in order to stop the rotation of the inertial body, the check valve 39A is biased by the valve spring 40A and moves to the valve seat.
41A, and the hydraulic pressure source side oil passage 21A and the actuator side oil passage
In the counterbalance valve 23, the pressure in the oil chamber 26A decreases, and the spool 25 is urged by the return spring 27B to return to the neutral position (a), thereby blocking communication between the hydraulic pressure source side oil passage 21B and the actuator side oil passage 22B.

However, the hydraulic motor 1 continues to rotate due to the inertial force of the inertial body connected to the output shaft 13, so even after the directional control valve 5 switches to the neutral position (a), it continues to suck in pressure oil from the primary actuator-side oil passage 22A and discharge it into the secondary actuator-side oil passage 22B, performing a pumping action. The pressure in the secondary actuator-side oil passage 22B gradually increases because the return of pressure oil to the tank 4 is prevented by the counterbalance valve 23 and the check valve 39B.

The pressure oil in the actuator side oil passage 22B flows through the throttle passage 3
4B into the valve guide 29B, and from the inside of the valve guide 29B through the throttle passage 37B, acts on the piston 36B.
2A. During this time, the valve guide 29B
The pressure in the valve guide 29B is kept lower than the pressure in the secondary actuator-side oil passage 22B due to the pipe loss in the throttle passages 34B and 37B, so the main valve element 32B opens at a pressure lower than the valve-opening set pressure _P0 and relieves at a low pressure. Next, when the piston 36B abuts against the valve casing 18 and stops, the pressure in the valve guide 29B becomes equal to the pressure in the actuator-side oil passage 22B,
The pressure rises to a predetermined valve opening setting and is relieved at high pressure.

While the relief valve 28B is open, the actuator-side oil passages 22A, 22B communicate with each other, forming a closed circuit made up of the actuator-side oil passages 22A, 22B and the supply/discharge passages 9, 10, etc. As a result, the pressure oil discharged from the hydraulic motor 1 is converted into thermal energy while passing through the relief valve 28B, and the inertial force of the inertial body is absorbed, generating a braking force.

The directional control valve 5 is set to a neutral position (a) and a switching position (c).
The same applies when switching between
The explanation will be omitted.

In the brake valve according to the above-mentioned prior art, the relief valves 28A and 28B with shockless function and the accumulators 38A and 38B are used.
The pressure oil is released at low pressure for a certain period of time (low pressure relief set time) until 6B stops sliding, and a sudden braking action is applied to prevent the shock caused by the sudden stopping of the inertial body.

However, in the prior art, even when the directional control valve 5 is switched from the neutral position (a) to the switching positions (b) and (c) to drive the hydraulic motor 1, the pressure oil discharged from the hydraulic pump 3 flows into each of the actuator-side oil passages 22A, 22B,
Since the pressure acts on the pistons 36A, 36B via the throttle passages 34A, 34B, etc., the driving pressure is kept low until the pistons 36A, 36B come into contact with the valve casing 18 and stop. This causes a problem that when the load is large, such as when climbing a hill or steering, the drive of the hydraulic motor 1 is delayed and responsiveness is reduced.

After a certain time (low pressure relief time) has elapsed until the pistons 36A and 36B stop sliding, the hydraulic motor 1
However, because the vehicle suddenly starts moving, the driver feels as if the inertial body is suddenly operating, which causes a problem of a significant decrease in safety.

Furthermore, since the low-pressure relief setting time of the accumulators 38A, 38B is determined by the flow path area of the throttle passages 34A, 34B and the throttle passages 37A, 37B formed in the main valve bodies 32A, 32B, when the ambient temperature drops, the viscosity of the pressure oil increases, making the low-pressure relief setting time longer than at room temperature, further worsening responsiveness and significantly reducing safety.

The present invention has been made in consideration of the problems of the prior art described above, and aims to provide a brake valve that can suppress the low-pressure relief operation of the relief valve when the actuator, which is an inertial body, is started, thereby improving responsiveness and safety.

DISCLOSURE OF THE INVENTION In order to solve the above-mentioned problems, a brake valve according to the present invention comprises a casing, a pair of oil passages that are provided within the casing and connected between a hydraulic power source side and an actuator side, a counterbalance valve that is located closer to the hydraulic power source side and provided midway through each of the passages, a pair of relief valves that are located closer to the actuator side and provided midway through each of the oil passages and are equipped with accumulators for maintaining a low valve opening pressure for a certain period of time, and a pair of check valves that are located between each relief valve and the counterbalance valve and are provided midway through each of the oil passages, allowing pressurized oil to flow from the hydraulic power source side toward the actuator side and preventing reverse flow.

The feature of the means adopted by the present invention is that the relief valves and check valves are arranged coaxially, and when each check valve is open, the inlet side of each relief valve is substantially closed.

In this case, it is preferable that each check valve has a notch formed on the side facing each relief valve.

It is also preferable that a cylindrical throttle member having a radially-drilled through hole on the side facing each relief valve is slidably inserted into each check valve, and that a valve spring is provided between each check valve and the throttle member to bias each throttle member toward each relief valve.

With the above-described configuration, when pressure oil discharged from the hydraulic source flows into the oil passage, the corresponding check valve opens, allowing the pressure oil to flow toward the actuator.

In this case, each relief valve and each check valve are arranged coaxially, and each check valve essentially closes the inlet of each relief valve when open, thereby restricting the flow of pressurized oil to the relief valve when the check valve is open, thereby preventing the relief valve from performing relief operation in a low-pressure state and improving responsiveness at startup.

BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a longitudinal sectional view of a brake valve according to a first embodiment of the present invention.

FIG. 2 is a cross-sectional view taken along the line II-II in FIG.

FIG. 3 is an enlarged cross-sectional view showing the main part of FIG.

FIG. 4 is a cross-sectional view of a brake valve according to a second embodiment of the present invention.

FIG. 5 is an enlarged cross-sectional view showing the main part of FIG.

FIG. 6 is a hydraulic circuit diagram provided with a brake valve according to the prior art.

FIG. 7 is a vertical cross-sectional view of the brake valve shown in FIG.

8 is a cross-sectional view taken along the line VIII-VIII in FIG.

FIG. 9 is an enlarged cross-sectional view showing the main part of FIG.

BEST MODE FOR CARRYING OUT THE INVENTION An embodiment of the present invention will be described below with reference to Figures 1 to 5. In the embodiment, the same components as those in the prior art shown in Figures 6 to 9 are designated by the same reference numerals, and their description will be omitted.

First, FIGS. 1 to 3 show a first embodiment of the present invention.

In the drawing, reference numeral 51 denotes a brake valve of this embodiment provided between the hydraulic motor 1 and the directional control valve 5. The brake valve 51 is provided integrally with the hydraulic motor 1 in a manner similar to the brake valve 6 described in the prior art.
The counterbalance valve 23 and the oil passages 53A and 53B described later are connected to each other.
When the directional control valve 5 is switched to the neutral position (a), one of the relief valves 59A, 59B opens to relieve the pressure oil on the high-pressure side of each of the pipes 2A, 2B to the low-pressure side, thereby applying a braking force to the hydraulic motor 1.

52 is provided to cover the opening side of the motor casing 7,
The valve casing 52 is firmly fixed by bolts 19, 19, . . . , and the valve casing 52 has oil passages 53A, 53B, as shown in FIG. 2, in a manner similar to the valve casing 18 described in the prior art.
are integrally formed, and a counterbalance valve 23, relief valves 59A, 59B, check valves 70A, 70B, etc. are provided.

Reference numerals 53A and 53B denote a pair of oil passages according to this embodiment that are provided in the valve casing 52 and form part of the pipes 2A and 2B.
Similar to the oil passages 20A, 20b described in the prior art, the oil passages 53A, 53B are generally composed of hydraulic source-side oil passages 54A, 54B connected to the hydraulic pump 3 via the directional control valve 5, and actuator-side oil passages 55A, 55B connected to the hydraulic motor 1 via the respective supply and discharge passages 9, 10, but inlet-side communicating passages 56A, 56B and outlet-side passages 57A, 57B, which will be described later, are provided midway through the respective actuator-side oil passages 55A, 55B.

Reference numerals 56A and 56B denote inlet-side communication passages provided midway through the actuator-side oil passages 55A and 55B, and the inlet-side communication passages 56A and 56B
The base ends thereof communicate with the actuator-side oil passages 55A and 55B, respectively, and the tip ends thereof communicate with the interiors of the valve seat members 62A and 62B, which will be described later.

Reference numerals 57A and 57B denote outflow side communicating passages serving as relief passages provided at the ends of the actuator side oil passages 55A and 55B, and the base ends of the outflow side communicating passages 57A and 57B communicate with the ends of the actuator side oil passages 55A and 55B, respectively, and the tip ends extend toward the opposing actuator side oil passages 55B and 55A, forming oil chambers 58A and 58B.

59A, 59B are located near the hydraulic motor 1 side and are connected to the valve casing 5
3, the relief valves 59A and 59B are connected to valve guides 60A and 60B, which will be described later, in a manner similar to the relief valves 28A and 28B described in the prior art.
0B, main valve bodies 63A, 63B, pistons 67A, 67B, etc., constitute a crossover relief valve with a shockless function. However, the relief valves 59A, 59B are provided on approximately the same axis as the hydraulic source-side oil passages 54A, 54B.

Reference numerals 60A and 60B denote valve guides disposed within the valve casing 52 and positioned approximately coaxially with the hydraulic pressure source-side oil passages 54A and 54B, with the distal ends of the valve guides 60A and 60B facing the interiors of the respective oil chambers 58B and 58A. Also, communication holes 61A and 61B are formed at the distal ends of the valve guides 60A and 60B.

Reference numerals 62A and 62B denote valve seat members positioned opposite the valve guides 60A and 60B and arranged approximately coaxially with the hydraulic pressure source side oil passages 54A and 54B. The base ends of the valve seat members 62A and 62B are in contact with the oil chambers 58.
The valve casing 52 is provided with its distal end facing into the inlet-side communication passages 56A, 56B.
The distal end side of each of the valve seat members 62A and 62B is located on the inlet side of the relief valves 59A and 59B.

63A, 63B are valve guides for seating on and separating from the valve seat members 62A, 62B, respectively.
The main valve bodies 63A, 63B are slidably mounted on the valve seats 60A, 60B, and the main valve bodies 63A, 63B are normally biased in the valve-closing direction by valve springs 64A, 64B. The main valve bodies 63A, 63B are also formed with throttle passages 65A, 65B in the axial direction. In this embodiment, the main valve bodies 63A, 63B are also provided with valve seat members 62A, 62B.
If the diameter of the seating portion with the valve guides 60A and 60B _{is d1} and the diameter of the sliding portion with the valve guides 60A and 60B is _d2 , then the relationship _d1 > _d2 holds, and the main valve bodies 63A and 63B open and close in the same way as in the prior art.

66A and 66B are the outer circumferential sides of the valve guides 60A and 60B and the valve casing 52
and a cylindrical oil chamber formed between the pistons 67A and 67B, the inner ends of which communicate with the oil chambers 58A and 58B, respectively. Pistons 67A and 67B are slidably provided in the oil chambers 66A and 66B, respectively. Valve guides 68A and 68B are valve guides 60A and 60B.
10 shows a throttle passage formed in the axially intermediate portion of each of the
The oil chambers 66A, 66B, the pistons 67A, 67B, the throttle passage 68A,
These and 68B constitute accumulators 69A, 68B which are substantially the same as the accumulators 38A, 38B described in the prior art.

70A and 70B are the counterbalance valve 23 and the relief valves 59A and 59B
The check valves 70A and 70B are located between the oil passages 53A and 53B, and are provided in the middle of the oil passages 53A and 53B.
Similar to the check valves 39A and 39B described in the prior art, valve seats 72A and 72B are formed between the hydraulic pressure source side oil passages 54A and 54B and the actuator side oil passages 55A and 55B by valve springs 71A and 71B.
However, like the relief valves 59A, 59B, the check valves 70A, 70B are mounted on substantially the same axis as the hydraulic-source-side oil passages 54A, 54B, and are formed with cutouts 73A, 73B as passages on the base ends facing the valve seat members 62A, 62B.

The brake valve according to this embodiment has the above-described structure, and its operation will now be described.

First, when the driver switches the directional control valve 5 from the neutral position (a) to the switching position (b), pressure oil from the hydraulic pump 3 flows into the hydraulic source side oil passage 54A via the pipe 2A, and the check valve 70A
The check valve 70A lifts off the valve seat 72A against the spring force of the valve spring 71A and opens. At the same time as the valve opens, the check valve 70A comes into contact with the tip side of the valve seat member 62A, and essentially blocks the passage between the inside of the valve seat member 62A, which is the inlet side of the relief valve 59A, and the inlet-side communicating passage 56A.

When the check valve 70A is opened, the hydraulic pressure in the hydraulic pressure source side oil passage 54A also acts on the inlet side communication passage 56A.
The inside of the valve seat member 62A is only communicated with through the notch 73A, and the inlet side communication passage 5 is closed by the throttling effect of the notch 73A.
The flow of pressure oil from 6A into the valve seat member 62A is restricted.

In this state, when the pressure oil in the valve seat member 62A acts on the piston 67A via the throttle passage 65A of the main valve element 63A and the throttle passage 68A of the accumulator 69A, the piston 67A slides in the oil chamber 66A and comes into contact with the oil chamber 58B side of the valve casing 52 and stops. As a result, the main valve element 63A of the relief valve 59A is set from the low pressure relief state to the high pressure relief state without actually opening for reasons that will be described later. Note that, regarding the accumulator 69B on the relief valve 59B side, the pressure oil in the actuator side oil passage 55A, which becomes high pressure, flows through the outlet side passage 58A.
7A and flows into the oil chamber 66B via the oil chamber 58A, displacing the piston 67B upward in the drawing and maintaining the main valve body 63B of the relief valve 59B in a state where it can relieve low pressure.

Thus, when the check valve 70A is open and in contact with the valve seat member 62A, the piston 67A of the accumulator 69A is in contact with the oil chamber
66A and stops when it comes into contact with the oil chamber 58B side of the valve casing 52, due to the throttling effect of the notch 73A.
Only a flow rate sufficient to allow the piston 67A to slide and displace is supplied. Therefore, the pressure inside the valve seat member 62A is maintained lower than the pressure in the inlet-side communicating passage 56A, and the spring force of the valve spring 64A keeps the main valve element 63A substantially closed, preventing low-pressure relief as in the prior art. Meanwhile, even if the main valve element 63A opens to relieve low pressure until the check valve 70A abuts against the valve seat member 62A, the pressure inside the valve seat member 62A drops below the pressure on the inlet-side communicating passage 56A side, causing the check valve 70A to be instantly attracted to the valve seat member 62A, immediately stopping low-pressure relief, and resulting in an extremely small flow rate to the low-pressure side. Therefore, the check valve 70A
When the valve opens against the spring force of the valve spring 71A, the pressure in the actuator side oil passage 55A is immediately maintained at a high pressure.

Next, the pressure oil in the high-pressure source side oil passage 54A flows through the actuator side oil passage 55A, the supply and discharge passages 9, etc., to the cylinder 11A of the rotor 11.
When the oil flows into the cylinder 11A, the piston 14 reciprocates within the cylinder 11A, causing the shaft 16 to slide along the cam surface 7A. The rotor 11 then rotates around the control pin 8, and the rotational force of the rotor 11 is output to the outside from the output shaft 13, rotating the inertial body.

In addition, when pressure oil flows into the hydraulic pressure source side oil passage 54A,
A part of this pressure oil flows through the throttle passage of the spool 25 into the oil chamber 26
2. As a result, the land 25B also moves rightward, connecting the hydraulic pressure source side oil passage 54B and the actuator side oil passage 55B, and the counterbalance valve 23 switches from the neutral position (a) to the switching position (b). Then, the pressure oil pushed out from the cylinder 11A by the compression stroke of the piston 14 flows into each supply and discharge passage 10.
etc. into the actuator side oil passage 55B, and is discharged from the actuator side oil passage 55B through the hydraulic source side oil passage 54B to the outside of the brake valve 6, and then returns to the tank 4 through the directional control valve 5 and the pipe 2B.

On the other hand, when the driver switches the directional control valve 5 from the switching position (b) to the neutral position (a) to stop the rotation of the inertial body, pressure oil is supplied to the check valve 70A via the notch 73A, so the check valve 70A is urged by the valve spring 71A to seat on the valve seat 72A, blocking the communication between the hydraulic power source side passage 54A and the actuator side passage 55A. Also, in the counterbalance valve 23, the pressure in the oil chamber 26A decreases, and the spool 25 is urged by the return spring 27B to return to the neutral position (a), thereby blocking the communication between the hydraulic power source side passage 54B and the actuator side passage 55B.

Then, the hydraulic motor 1 is forced to rotate due to its inertial force,
When pressure oil is discharged into the actuator-side oil passage 55B, the pressure oil in the actuator-side oil passage 55B flows from the inlet-side communication passage 56B through the throttle passage 65B into the valve guide 60B, and acts on the piston 67B from the valve guide 60B through the throttle passage 68B.
During this time, the pressure inside the valve guide 60B is kept lower than the pressure inside the actuator-side oil passage 55B due to the line loss of the throttle passages 65B and 68B, so the main valve element 63B opens at a pressure lower than the predetermined valve opening set pressure and relieves at a low pressure.
Next, when the piston 67B comes into contact with the valve casing 52 and stops, the pressure in the valve guide 60B flows through the actuator-side oil passage 55B.
The pressure inside the valve rises to the specified valve opening pressure and is relieved at high pressure.

While the relief valve 59B is open, the actuator-side oil passages 55A and 55B communicate with each other via the outflow-side communication passage 57A, and the actuator-side oil passages 55A and 55B communicate with the supply and discharge passages 9 and 1.
0, etc., and the pressure of the pressure oil discharged from the hydraulic motor 1 is converted into thermal energy while passing through the relief valve 59B, generating a braking force.

The directional control valve 5 is set to a neutral position (a) and a switching position (c).
The same applies to the case where the switch is made between the above, and therefore the explanation will be omitted.

Thus, according to this embodiment, the relief valves 59A, 59B and the check valves 70A, 70B are disposed approximately coaxially with the hydraulic source oil passages 54A, 54B, and when the check valves 70A, 70B are opened,
Since the check valves 70A, 70B are configured to substantially close the tip ends of the valve seat members 62A, 62B, which are the inlet side of the relief valves 59A, 69B, pressure oil is supplied into the valve seat members 62A, 62B only through the notches 73A, 73B formed in the check valves 70A, 70B.
The pistons 67A, 67B of the valve seats 62A, 62B slide in the oil chambers 66A, 66B until they come into contact with the valve casing 52.
The pressure inside B becomes low, and the main valve bodies 63A, 63B of the relief valves 59A, 59B are substantially kept closed.
When the inertial body starts to open the check valves 70A and 70B, the leaf valves 59A and 59B can be prevented and suppressed from relieving at a low pressure.

As a result, pressure oil can be supplied quickly to the hydraulic motor 1 to drive it to rotate, improving responsiveness, and reliably preventing the hydraulic motor 1 from suddenly rotating after the low-pressure relief setting time has elapsed, as occurs with relief valves in the prior art, thereby improving safety and
Reliability can be significantly improved. In addition, since the relief valves 59A, 59B and the check valves 70A, 70B are arranged coaxially, the relief valves 59A, 59B and the check valves 70A, 70B can be efficiently accommodated within the valve casing 52, making it possible to form the brake valve 51 as a whole compact. Furthermore, since the cutouts 73A, 73B are formed in the check valves 70A, 70B,
When the check valves 70A and 70B are open,
and the valve seat members 62A, 62B, pressure oil is interposed between the check valves 70A, 80 and the valve seat members 62A, 62B, thereby preventing the occurrence of adhesion. When the directional control valve 5 is switched to the neutral position (a),
B can be closed immediately.

Next, FIGS. 4 and 5 show a second embodiment of the present invention.
The feature of this embodiment is that a throttle member is provided between the valve seat member and the check valve. In this embodiment, the same components as those in the first embodiment are designated by the same reference numerals, and their description will be omitted.

In the drawing, 81A and 81B indicate check valves according to this embodiment provided in the oil passages 53A and 53B, and the check valves 81A and 81B
Although the check valves 81A and 81B are formed in substantially the same manner as the check valves 70A and 70B described in the first embodiment,
The outer peripheries of throttle members 82A and 82B, which will be described later, are slidably inserted into the grooves.

Reference numerals 82A and 82B denote cylindrical throttle members that are located between the valve seat members 62A and 62B and the check valves 81A and 81B and are slidably provided within the check valves 81A and 81B.
The valve seat member 62 includes stoppers 83A and 83B formed on the tip end side and abutting against the check valves 81A and 81B, spring bearings 84A and 84B formed integrally with the valve seat member 62 and projecting radially inward from the inner circumferential side of the intermediate portion, and a valve seat member 62.
The throttle members 82A, 82B are generally composed of a plurality of communicating holes 85A, 85B bored radially at the base ends thereof facing the valve seats 62A, 62B. Valve springs 86A, 86B that constantly urge the check valves 81A, 81B toward the valve seats 72A, 72B are provided between the spring bearings 84A, 84B and the check valves 81A, 84B. The throttle members 82A, 82B are urged by the spring force of the valve springs 86A, 86B, with their base ends pressed against the inner circumferential tips of the valve seat members 62A, 62B.

When the check valves 81A, 81B are open, the throttle members 82A, 82B throttle and allow the pressure oil to flow between the valve seat members 62A, 62B and the check valves 81A, 81B, thereby reducing the pressure of the relief valve 59.
This suppresses the low pressure relief operation of the check valves 81A and 81B, and also quickly closes the check valves 81A and 81B when the hydraulic motor 1 is stopped.

Furthermore, the throttle members 82A, 82B generate a pressure difference between the inner and outer peripheries thereof through the communication holes 85A, 85B, thereby preventing chattering when the check valves 81A, 81B are opened, and by guiding the check valves 81A, 81B from the inner periphery side, the check valves 81A, 81B are guided to the valve seat member 62.
A, 62B is seated coaxially on the tip side.

Thus, even in this embodiment configured as above, it is possible to obtain substantially the same effects as in the first embodiment. That is, when the check valves 81A, 81B are opened, the communication holes 85A, 85B of the throttle members 82A, 82B are opened.
As a result, similar to the first embodiment, the pistons 67A and 67B of the accumulators 69A and 69B are forced to move in the oil chambers 66A and 66B.
The main valve bodies 63A, 63B of the relief valves 59A, 59B remain closed until the high pressure relief state is reached.
When the inertial body is activated, each of the relief valves 59A and 59B can be prevented from relieving low pressure.

However, in this embodiment, the throttle members 82A, 82B are slidably provided between the valve seat members 62A, 62B and the check valves 81A, 81B, and the communication holes 85A, 85B are formed in the throttle members 82A, 82B.
This creates a large pressure difference between the inside and outside of the check valves 81A and 81B.
Furthermore, by guiding the check valves 81A and 81B from the inner circumferential side, the throttle members 82A and 82B can be positioned and moved coaxially with the valve seat members 62A and 62B.

In the above embodiments, the accumulators 69A and 69B are provided on the outer periphery of the valve guides 60A and 60B.
Alternatively, the accumulator may be located within the valve guide or may be located entirely remotely.

In addition, in each of the above embodiments, the radial piston type hydraulic motor 1 has been described as an example of the actuator.
The present invention is not limited to this, and other actuators such as a swash plate type hydraulic rotary machine and a bent axis type hydraulic rotary machine may be used.

Furthermore, the present invention is not limited to hydraulic excavators, hydraulic cranes, etc., but can be widely applied to brake valves of tracked vehicles that are driven by hydraulic motors, construction machines that have a lower traveling body and an upper rotating body, etc.

Furthermore, in the first embodiment, the accumulators 69A and 69B
The pistons 67A and 67B slide and displace, and the relief valves 59A and 59B
During the period from the low pressure relief state to the high pressure relief state, the passage between each relief valve 59A, 59B and the check valve 70A, 70B is substantially closed, and only the notched portions 73A, 73B are formed. Therefore, even if the main valve bodies 63A, 63B of the relief valves 59A, 59B move slightly toward the valve opening side and leak into the tank during low pressure relief, the passage between each relief valve 59A, 59B and the check valve 70A, 70B is substantially closed, and only the notched portions 73A, 73B are formed.
The flow rate to the low-pressure side is extremely small due to the throttling effect of 73A and 73B, which means that there is no effect on the responsiveness of the hydraulic motor 1 during orbit, and the hydraulic motor 1 can be started quickly. This also applies to the second embodiment.

INDUSTRIAL APPLICABILITY As described above in detail, according to the present invention, each relief valve and check valve are arranged coaxially, and when each check valve is open, the inlet side of each relief valve is closed.
When the corresponding check valve is opened by pressure oil from the hydraulic source, the relief valve can be maintained in an open state. As a result, pressure oil can be quickly supplied to the actuator to drive it, improving responsiveness. Unlike the prior art, after the low-pressure relief setting time of the accumulator has elapsed,
This reliably prevents the actuator from suddenly operating, improving safety and reliability. Moreover, because the relief valves and check valves are arranged coaxially, they can be efficiently housed within the casing, allowing the entire brake valve to be made compact.

Claims (3)

[Claims] [Claim 1] A brake valve comprising: a casing; a pair of oil passages provided within the casing and connected between a hydraulic power source side and an actuator side; a counterbalance valve located closer to the hydraulic power source side and provided midway through each of the oil passages; a pair of relief valves located closer to the actuator side and provided midway through each of the oil passages and equipped with accumulators for maintaining a low valve opening pressure for a certain period of time; and a pair of check valves located between each of the relief valves and the counterbalance valve and provided midway through each of the oil passages, allowing pressurized oil to flow from the hydraulic power source side to the actuator side and preventing reverse flow, wherein each of the relief valves and each of the check valves are arranged coaxially and each of the check valves is configured to substantially close the inlet side of each of the relief valves when open. 2. A brake valve according to claim 1, wherein each of said check valves has a notch formed on the side facing each of said relief valves. [Claim 3] A brake valve as described in claim 1, wherein a cylindrical throttling member having a communicating hole drilled radially on the side facing each relief valve is slidably inserted into each check valve, and a valve spring is provided between each check valve and the throttling member to urge each throttling member toward each relief valve.